1. Field of the Invention
This invention relates to rotary screw compressors, and more particularly to a compressor and a method of operation that will provide automatic compensation against axial thrust forces imposed on the compressor rotor bearings.
2. Description of the Prior Art
Rotary screw compressors comprise a housing with working fluid inlet and outlets, rotor bores and a rotor assembly mounted on bearings for rotation in the rotor bores. The rotor may comprise a single rotor or male and female screw-type rotors having intermeshed lands and grooves. Rotation of the rotor causes a working fluid to be taken from the low pressure inlet or suction side, and gradually compressed in chambers created by the lands and grooves. The high pressure fluid is then discharged through the high pressure outlet.
The capacity of the compressor and the volume ratio of the compressor, sometimes called compression ratio, are controlled by various types of valve arrangements. One type of valve arrangement used to regulate the capacity and volume ratio is termed a slide valve. If a slide valve is used, the compressor housing is provided with a slide valve receiving recess which connects the rotor bores in fluid communication with the low pressure inlet. The slide valve is mounted and operative to either close this recess or open it thereby providing a variable size bypass opening to bypass some compressed fluid back to this inlet to control the compressor capacity.
The volume ratio of the compressor depends upon the period of time fluid remains trapped in the rotor chambers. As the rotors rotate, the rotor chambers become progressively smaller which reduces the volume of the fluid therein and increases its pressure. Therefore, the longer the period of time that fluid remains trapped in the rotor chambers, the smaller its volume becomes. The slide valve is adjustable to regulate the period of time fluid is trapped in the rotor chambers and increasing or decreasing retention time increases or decreases the compressor volume ratio.
An inherent differential pressure .DELTA.P exists between the low pressure inlet and the high pressure outlet sides of the compressor. This .DELTA.P pressure acts against the end faces of the rotors and generates axial thrust forces tending to move the rotors toward the low pressure inlet side. These axial thrust forces must be absorbed by the bearings and such forces can generate extremely high axial bearing loads which overload the bearings under many normal operating parameters. However, under other operating parameters very little or no axial thrust force may be generated with the consequence that the bearings are substantially under loaded.
It has long been known that high axial bearing loads produce greater friction and higher operating temperatures on the thrust bearings which greatly reduce their operating life. For example, at face-to-face bearing loads of 10,000 lbs, the bearing life will be under 2000 hours, or less than three months. Replacement of these bearings is extremely expensive in bearing cost, labor cost, and compressor downtime. It has also been known that to guarantee satisfactory performance of both roller and ball bearings, they must always be subject to a given minimum load especially if they run at high speeds such as in compressors where the inertia forces of the bearing elements and cage, and friction in the lubricant, may cause damaging sliding movements to occur between the bearing elements and their raceways. Therefore, both the absence of a minimum load and the presence of a high axial bearing load can damage and drastically shorten the bearing life.
The problem of a short bearing service life in compressors has been recognized for decades and many solutions to solve it have been suggested. The prior art teaches that the high axial thrust forces should be opposed by a counterbalancing force acting in the opposite direction. To accomplish this, U.S. Pat. No. 3,161,349 issued Dec. 15, 1964 to L. B. Schibbye teaches that a counterbalancing piston should be mounted on the rotor in a compartment that is connected to a source of pressurized compressor lubricating oil provided by a pump driven by the compressor. The lubricating oil pressure, in function, reflects the discharge pressure of the compressor and thus generates a counterbalancing force which is a function of the differential pressure .DELTA.P of the compressor. This counterbalancing piston will exert a force on the bearing that is counter to the axial thrust force. However, as shown in FIG. 4, developing a force that references discharge pressure produces a force WDPT which is a straight line over the output capacity of the compressor as indicated by the 0-100 psia range of suction pressures shown.
Refrigeration and air conditioning compressors are equipped with some type of valve arrangement as previously discussed for varying the capacity of the compressor between maximum and minimum levels. The axial thrust force on the rotor will vary as the capacity of the compressor varies. The resulting axial bearing load at a minimum capacity will be about one-half of the axial bearing load that exists at a maximum capacity. Because, as discussed above, a bearing must always have a minimum loading to prevent failure, a dilemma always exists between two design parameters. First, for long bearing life a counterbalance force applying piston must be sized (areawise) to be as large as possible to offset as much of the axial thrust force as possible at maximum capacity. Second, for long bearing life a counterbalance force applying piston must be sized small enough to prevent overbalancing against the axial thrust force at minimum capacity to prevent underloading the bearing. Therefore, if one sizes the counterbalancing piston to meet the second parameter, there is not enough counterbalancing force at maximum capacity and the bearing life is shortened. If one sizes the counterbalancing piston to meet the first parameter, the bearings will be unloaded at certain minimum capacity conditions and the bearing life is shortened because the required minimum bearing load is not maintained.
This dilemma is illustrated in FIG. 4. Plot FW/OCB (force without counterbalancing) shows that during operation the force varies at maximum capacity from approximately 3920 to 9800 lbs at a constant .DELTA.P of 100 psi. If one references discharge pressure for counterbalancing, the force WDPT available for counterbalancing is approximately 1335 lbs for a typically sized counterbalancing piston for a particular size rotor no matter what the suction pressure is as long as the .DELTA.P is constant. Therefore, at maximum capacity and 10 psi suction pressure (WR1) the net axial force FDPT-1 available for counterbalancing would be 4400-1335=3065 lbs. The bearing load resulting from this force would result in acceptable bearing life. However, at minimum compressor capacity (FIG. 5), the axial force without counterbalancing would be as shown at WR1 in FIG. 5 and the net bearing load FDPT-1 would be 2200-1335=895 lbs. This loading is far below the bearing manufacturer's recommended minimum load of 2000 lbs and will result in unacceptable bearing life. Referring back again to maximum capacity (FIG. 4), at a 90 psia suction pressure (WR2), the net axial force FDPT-2 would be 9100-1335=7765 lbs. This allows a bearing load that is far too high and would result in a bearing life of less than one year. At minimum compressor capacity (FIG. 5) at 90 psia (WR2), the net bearing load FDPT-2 (from FIG. 5) would be 4550-1335=3215 lbs. This would be an acceptable minimum bearing load.
The following is Table 1 which lists typical values of relevant operating parameters of a compressor of conventional prior art design at .DELTA.P=100 psi wherein discharge pressure of the compressor is sensed and used to provide a pressure for application to a counterbalancing piston. These typical values are for a particular size of standard compressor, balance piston, and bearing arrangement.
TABLE 1 ______________________________________ .DELTA.P = 100 psi Prior Art (Discharge Pressure) ______________________________________ Suction Pressure 10 10 90 90 Compressor Capacity Min Max Min Max Axial Force 2200 4400 4550 9100 FW/OCB Counterbalance Force 1335 1335 1335 1335 WDPT Net Bearing Load 895 3065 3215 7765 FW/OCB - WDPT ______________________________________
There have been many arrangements suggested by the prior art to reduce the adverse effects of these problems. U.S. Pat. No. 3,388,854 issued Jun. 18, 1968 to Olofsson et al uses a spring 35 acting on the thrust bearings. This spring exerts axial thrust on the rotor in the opposite direction to the axial force exerted by the thrust counterbalancing piston to distribute axial thrust more evenly.
U.S. Pat. No. 3,811,805 issued May 21, 1974 to Moody, Jr. et al recognizes that the thrust balance pistons can exert a counterbalancing force that overcompensates for the axial thrust forces. Moody, Jr. et al states that the adverse effects can be overcome by providing a hydrodynamic fluid bearing between the end faces of both female and male screws and a fixed thrust surface of the housing. An oil film is maintained between these two components to reduce wear but this does not fully address the problem of overloading or underloading the bearings.
U.S. Pat. No. 4,180,089 issued Dec. 25, 1979 to Webb also correlates the biasing of the thrust balance pistons to the discharge pressure of the compressor. Webb uses a valve structure in the high pressure lubrication oil line to attenuate the pressure applied to the thrust balance piston so that it will be approximately 20 psi below whatever the compressor discharge pressure is. However, the basic problem of overloading and underloading is not solved.
U.S. Pat. No. Reissue 32,055 issued Dec. 24, 1985 to Schibbye et al discloses that high pressure lubricating oil should be supplied to the thrust balance piston on the low pressure end of the male rotor; that a mean lubricating oil pressure should be applied to the high pressure ends of both the male and female rotors; and that an axial connection passage be provided from the high pressure end of the female rotor to the female rotor balancing piston at the low pressure end thereof to keep both ends at the mean pressure. Thus, the low pressure end of the male rotor is at a high thrust balancing pressure and the low pressure end of the female rotor is at a lower mean thrust balancing pressure to help increase service life of the bearings but does not fully address the problem of underloading and overloading the bearings.
U.S. Pat. No. 4,964,790 issued Oct. 23, 1990 to Scott states that in the prior art "the balancing pressure on the pistons is not responsive to the various operative parameters other than outlet pressure of the rotary compressor." Scott discloses a complex system using a microprocessor control for computing a net counterbalancing force in response to inputs or sensed parameters relating to the pressure of gas at the inlet and outlet of the compressor, and regulates a variable valve of an oil pump responsive to the microprocessor signal to control the amount of thrust balancing oil pressure applied to the counterbalancing pistons.
All of the thrust balancing systems of the prior art are either unduly complex in construction and function and therefore expensive to manufacture and service, or do not supply a counterbalancing force which correlates the axial bearing load through the full range of suction pressures existing between a minimum and maximum compressor working range as illustrated by plots WR1and WR2 of FIGS. 4 and 5.
Therefore, what is needed is a compressor having a simple, reliable, low cost thrust bearing force compensation arrangement and a method for its operation to produce a counterbalancing force correlated to the axial force on the rotors.